This application claims the priority of German application 198 38 754.7, filed in Germany on Aug. 26, 1998, the disclosure of which is expressly incorporated by reference herein.
The invention relates to an exhaust gas turbocharger for an internal-combustion engine whose exhaust gas turbine has a rotor disk with at least one semiaxial and one radial flow inlet cross-section and is equipped with a variable turbine geometry for the variable adjustment of at least one flow inlet cross-section, and having a compressor in the intake system of the internal-combustion engine, the ratio (D.sub.v /D.sub.T,m) of the compressor outlet diameter (D.sub.v) to the average turbine inlet diameter (D.sub.T,m) according to the adjustment of the variable turbine geometry being variably adjustable between a minimal value and a maximal value.
From German Patent Document DE 43 30 487 C1, an exhaust gas turbocharger for a combustion engine is known whose exhaust gas turbine has a rotor disk with a semiaxial and a radial flow inlet cross-section. In the area of the semi-axial flow inlet cross-section, rigid guide baffles are arranged, but in the area of the radial flow inlet cross-section, guide baffles with adjustable guide blades are arranged by way of which the surface of the radial flow inlet cross-section can be varied as a function of the guide blade position.
In the starting operation, the radial guide baffles are changed into the shut-off position so that the radial inflow to the rotor disk is blocked and the exhaust gas current flows through the semiaxial flow inlet cross-section to the rotor disk and impacts at a fluidically favorable angle onto the rotor disk. In the partial load range and full load range, the radial guide baffles are opened up so that a higher fraction of the exhaust gas current flows through the radial flow inlet cross-section.
The variable turbine geometry, which is implemented by way of the adjustable radial guide baffles, can be used in the engine braking operation for increasing the braking performance. For this purpose the guide blades are changed into the ram position which results in the buildup of a high exhaust gas back pressure. At the same time, exhaust gas flows at a high flow rate through the ducts between the guide blades or through the semiaxial flow cross-section and acts upon the turbine wheel, whereby the compressor output is increased and the combustion air supplied to the engine is subjected to excess pressure by the compressor.
As the result, the cylinder is acted upon with increased supercharging pressure on the inlet side. On the outlet side, an excess pressure exists between the cylinder outlet and the exhaust gas turbocharger which counteracts the blowing-off of the air compressed in the cylinder by way of decompression valves into the exhaust gas pipe system. In the engine braking operation, the piston must carry out compression work in the compression and push-out stroke against the high excess pressure in the exhaust gas pipe system, whereby a strong braking effect is achieved.
With respect to additional prior art, reference is made to German Patent Document DE 196 15 237 A1, International Patent Document WO 84/04136 and German Patent Document DE 197 27 140 C1 which was published later.
The invention is based on the problem of improving the efficiency of an exhaust gas turbocharger of the above-mentioned type in a broad operating range.
According to the invention, this problem is solved by providing an arrangement of the above-noted type wherein wherein the minimal value is lower than 1.1 and the average turbine inlet diameter (D.sub.T,m) is calculated according to the equation
D.sub.T,m =D.sub.T,x +(D.sub.T,r -D.sub.T,x)*A.sub.T,r /(A.sub.T,r +A.sub.T,x)
wherein
D.sub.T,x is the semiaxial turbine inlet diameter PA1 D.sub.T,r is the radial turbine inlet diameter PA1 A.sub.T,r is the area of the radial flow inlet cross-section PA1 A.sub.T,x is the area of the semiaxial flow inlet cross-section,
and wherein the ratio of the semiaxial flow inlet cross-section area to the radial flow inlet cross-section area when the semiaxial and radial flow paths are fully open amounts to approximately 0.1.
With respect to the fired operation of exhaust gas turbines, it must be taken into account that a large turbine wheel diameter results in efficiency advantages in the upper engine rotational speed range. Possible disadvantages of large turbine wheel diameters with respect to the fuel consumption, which may be the result of the higher inertia of the rotor disk, are overcompensated by the retaining possibility of the variable guide baffles. It was found that, at a ratio of the compressor wheel diameter to the turbine wheel diameter of maximally 1.1, the advantages of a large turbine wheel diameter prevail by far.
For calculating the ratio of the compressor wheel to the turbine wheel, an average value is first determined for the turbine wheel diameter. With a different weighting, the semiaxial turbine inlet diameter and the radial turbine inlet diameter are included in the calculation of the average value, for the weighting of these two components, the ratio of the surface of the radial flow inlet cross-section to the sum of the semiaxial and radial flow inlet cross-section being used. The semiaxial or radial flow inlet cross-section corresponds to the guide baffle outlet surface in the semiaxial or radial inflow. The average turbine wheel diameter is then used as the basis for calculating the ratio of the compressor wheel to the turbine wheel.
The maximal value of the ratio of the compressor wheel to the turbine wheel is reached when the average turbine wheel diameter becomes minimal. This is the case if the larger fraction of the exhaust gas flow rate is guided through the inflow with the smaller radial distance from the turbine axis--as a rule, the semiaxial inflow--. Inversely, the minimal value of the ratio of the compressor wheel to the turbine wheel is reached when the average turbine wheel diameter becomes maximal, which normally occurs in the case of a preferred radial inflow.
The average value of the turbine wheel diameter is no constant value but, because of the variable weighting factors, may vary according to the position of the variable turbine geometry between a maximal value and a minimal value. Accordingly, by way of the variable turbine geometry, the average value of the turbine wheel diameter can be set. The minimal value of the ratio of the compressor wheel to the turbine wheel, according to the invention, must not exceed the 1.1 value which means that the turbine wheel diameter must have a minimal value in relationship to the compressor wheel diameter.
As a result, the turbine of an exhaust gas turbocharger having a variably adjustable geometry can be optimally adjusted according to the efficiency improvement and consumption reduction criteria. Because of the variability of the turbine wheel diameter, instead of a single operating point, an operating range can be covered now which, in the case of a low load and rotational speed, is displaced in the direction of a more favorable fuel consumption and, in the partial load and full load range, is in the range of an optimal efficiency.
Because of the variability of the average turbine wheel diameter, this may also be called a quasi-variable turbine wheel which has the same characteristics as a turbine wheel whose diameter can be physically variably adjusted but which is not equipped with the complicated geometry and kinematics of a variable turbine wheel. The average turbine wheel diameter has the significance of a hypothetical diameter whose variation takes place by a semiaxial or radial approach flow of different intensity.
The calculation of the average value of the turbine wheel diameter is independent of the selection of the variable turbine geometry. When radial guide baffles with adjustable guide blades are used and when an axial slider is used in the semiaxial turbine inlet or when both above-mentioned shut-off devices are used, the relationship for the average value of the turbine wheel diameter can be used and can be the basis of the determination of the ratio of the compressor wheel diameter to the turbine wheel diameter. Optionally, a flap turbine can also be used.
In the case of a preferably selected minimal value of at least 0.9, but lower than 1.0, and a maximal value of 1.2, the ratio of the compressor wheel diameter to the turbine wheel diameter can be varied in a bandwidth of between at least 0.9 and maximally 1.2. This bandwidth sweeps over the efficiency maximum of the turbine.
If the minimal value is set to lower than 1.0, the hypothetical turbine wheel diameter in the position of the shut-off device of the variable turbine geometry which corresponds to this minimal value is larger than the compressor wheel diameter. The average diameter of the turbine wheel exceeds the diameter of the compressor wheel.
By way of the shut-off device of the variable turbine geometry, the exhaust gas flow rates to the turbine rotor disk can be adjusted. According to the shut-off device and the position of the shut-off device, differently large flow rates flow by way of the semiaxial inflow and the radial inflow to the rotor disk. The flow rates can additionally be influenced by the ratio of the semiaxial and the radial flow inlet cross-sections. The ratio of 0.1 of the semiaxial to the radial flow inlet cross-section when the guide baffles are open was found to be advantageous because, at this ratio, the semiaxial inflow has a narrow cross-section and, even at low exhaust gas back pressures, relatively high flow velocities can be reached in the semiaxial ring nozzle.
If radial guide baffles with adjustable guide blades are used in the area of the radial inflow as variable turbine geometry, the maximal value of the ratio of the compressor wheel to the turbine wheel is expediently reached when the guide baffles are closed; the minimal value is reached when the guide baffles are open.
The same applies to the use of an axial slider in the area of the radial inflow.
However, in another advantageous embodiment, a shut-off device can also be used in the area of the semiaxial inflow, in this case, the maximal value being reached when the shut-off device is open and the minimal value being reached when the shut-off device is closed.
Optionally, shut-off devices may be inserted in the semiaxial inflow as well as in the radial inflow.
In another solution according to the invention, an average wheel inlet angle is determined from a different weighting of the semiaxial and radial wheel inlet angle with respect to the turbine rotor disk. For the optimal adjustment for the semiaxial and the radial, in each case, ring-nozzle-shaped inflow, a range for the average wheel inlet angle of approximately 20 to 90.degree. is defined. The wheel inlet angle represents another degree of freedom which, in addition to the ratio of the compressor wheel to the turbine wheel, can be taken into account during the dimensioning.